EP0297682B1 - Hydraulisches Antriebssystem - Google Patents

Hydraulisches Antriebssystem Download PDF

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Publication number
EP0297682B1
EP0297682B1 EP19880201351 EP88201351A EP0297682B1 EP 0297682 B1 EP0297682 B1 EP 0297682B1 EP 19880201351 EP19880201351 EP 19880201351 EP 88201351 A EP88201351 A EP 88201351A EP 0297682 B1 EP0297682 B1 EP 0297682B1
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EP
European Patent Office
Prior art keywords
pressure
valve
hydraulic
pilot
control
Prior art date
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Expired - Lifetime
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EP19880201351
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English (en)
French (fr)
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EP0297682A2 (de
EP0297682A3 (en
Inventor
Eiki Izumi
Toichi Hirata
Yusaku Nozawa
Masahiko Shimotori
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Hitachi Construction Machinery Co Ltd
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Hitachi Construction Machinery Co Ltd
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Publication of EP0297682A3 publication Critical patent/EP0297682A3/en
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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B9/00—Servomotors with follow-up action, e.g. obtained by feed-back control, i.e. in which the position of the actuated member conforms with that of the controlling member
    • F15B9/02—Servomotors with follow-up action, e.g. obtained by feed-back control, i.e. in which the position of the actuated member conforms with that of the controlling member with servomotors of the reciprocatable or oscillatable type
    • F15B9/08—Servomotors with follow-up action, e.g. obtained by feed-back control, i.e. in which the position of the actuated member conforms with that of the controlling member with servomotors of the reciprocatable or oscillatable type controlled by valves affecting the fluid feed or the fluid outlet of the servomotor
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B13/00—Details of servomotor systems ; Valves for servomotor systems
    • F15B13/02—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors
    • F15B13/04—Fluid distribution or supply devices characterised by their adaptation to the control of servomotors for use with a single servomotor
    • F15B13/0401—Valve members; Fluid interconnections therefor
    • F15B13/0405—Valve members; Fluid interconnections therefor for seat valves, i.e. poppet valves
    • E—FIXED CONSTRUCTIONS
    • E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
    • E02F—DREDGING; SOIL-SHIFTING
    • E02F9/00—Component parts of dredgers or soil-shifting machines, not restricted to one of the kinds covered by groups E02F3/00 - E02F7/00
    • E02F9/20—Drives; Control devices
    • E02F9/22—Hydraulic or pneumatic drives
    • E02F9/2221—Control of flow rate; Load sensing arrangements
    • E02F9/2225—Control of flow rate; Load sensing arrangements using pressure-compensating valves
    • E—FIXED CONSTRUCTIONS
    • E02—HYDRAULIC ENGINEERING; FOUNDATIONS; SOIL SHIFTING
    • E02F—DREDGING; SOIL-SHIFTING
    • E02F9/00—Component parts of dredgers or soil-shifting machines, not restricted to one of the kinds covered by groups E02F3/00 - E02F7/00
    • E02F9/20—Drives; Control devices
    • E02F9/22—Hydraulic or pneumatic drives
    • E02F9/2221—Control of flow rate; Load sensing arrangements
    • E02F9/2232—Control of flow rate; Load sensing arrangements using one or more variable displacement pumps
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F15—FLUID-PRESSURE ACTUATORS; HYDRAULICS OR PNEUMATICS IN GENERAL
    • F15B—SYSTEMS ACTING BY MEANS OF FLUIDS IN GENERAL; FLUID-PRESSURE ACTUATORS, e.g. SERVOMOTORS; DETAILS OF FLUID-PRESSURE SYSTEMS, NOT OTHERWISE PROVIDED FOR
    • F15B11/00—Servomotor systems without provision for follow-up action; Circuits therefor
    • F15B11/16—Servomotor systems without provision for follow-up action; Circuits therefor with two or more servomotors
    • Y—GENERAL TAGGING OF NEW TECHNOLOGICAL DEVELOPMENTS; GENERAL TAGGING OF CROSS-SECTIONAL TECHNOLOGIES SPANNING OVER SEVERAL SECTIONS OF THE IPC; TECHNICAL SUBJECTS COVERED BY FORMER USPC CROSS-REFERENCE ART COLLECTIONS [XRACs] AND DIGESTS
    • Y10—TECHNICAL SUBJECTS COVERED BY FORMER USPC
    • Y10T—TECHNICAL SUBJECTS COVERED BY FORMER US CLASSIFICATION
    • Y10T137/00—Fluid handling
    • Y10T137/8593—Systems
    • Y10T137/87169—Supply and exhaust
    • Y10T137/87193—Pilot-actuated

Definitions

  • the present invention relates to a hydraulic drive system for hydraulic construction machines, such as hydraulic excavators and hydraulic cranes, the hydraulic drive system comprising: at least one hydraulic pump, at least first and second hydraulic actuators connected to said hydraulic pump through respective main circuits and driven by hydraulic fluid delivered from said hydraulic pump; first and second flow control valve means connected to said respective main circuits between said hydraulic pump and said first and second hydraulic actuators, pump control means for controlling a delivery pressure of said hydraulic pump; each of said first and second flow control valve means comprising first valve means having an opening degree variable in response to the operated amount of operation means, and second valve means connected in series with said first valve means for controlling a differential pressure between the inlet pressure and the outlet pressure of said first valve means; and control means associated with each of said first and second flow control valve means for causing said second valve means to control the differential pressure between the inlet pressur and the outlet pressure of said first valve means based on the inlet pressure and the outlet pressure of said first valve means, the delivery pressure of said hydraulic pump, and the maximum load pressure between said first and second hydraulic actuators.
  • Such hydraulic drive system is known from US-A-4,617,854, disclosing a system of the type wherein an auxiliary valve is disposed in the main circuit upstream of each flow control valve, the inlet and outlet pressures of the flow control valve are both introduced to a first one of opposite operating parts of the auxiliary valve, the delivery pressure of the hydraulic pump and the maximum load pressure among a plurality of hydraulic actuators are both introduced to a second one of the opposite operating parts thereof, and a pump regulator of the load sensing type is disposed which serves to hold the delivery pressure of the hydraulic pump a predetermined value above that maximum load pressure.
  • the load pressure of the flow control valve is compensated as known in the art. Also, by introducing the delivery pressure of the hydraulic pump regulated by the pump regulator and the maximum load pressure among the plurality of hydraulic actuators to a second one of the opposite operating parts of the auxiliary valve, in the combined operation of the plurality of hydraulic actuators having respective load pressures different from each other, it is made possible that even if the total of commanded flow rates (required flow rates) of the respective hydraulic actuators exceeds a maximum delivery flow rate of the hydraulic pump, the delivery rate of the hydraulic pump is distributed in accordance with relative ratios of the commanded flow rates to thereby ensure that hydraulic fluid is reliably passed to the hydraulic actuators on the side of higher load pressure as well.
  • the flow control valve and the auxiliary valve each comprise a spool valve which is relatively large in size, as they are both disposed in the main circuit. Accordingly, if the hydraulic circuit is subject to a higher pressure for energy saving, this would give rise to the problem of causing appreciable fluid leakage from those spool valves. Also, since the auxiliary valve is disposed in the main circuit through which a large flow rate passes, the system suffers from another problem, being an increasing pressure loss at the auxiliary valve.
  • each hydraulic actuator in the hydraulic drive system preferably should be supplied with a corresponding flow rate free of any effects from self-load pressure and respective load pressures of other hydraulic actuators.
  • hydraulic drive systems employed in construction machines such as hydraulic excavators to be affected by load pressures of any other hydraulic actuators or self-load pressure depending on the types of working members and the working modes thereof to be driven by the relevant hydraulic actuator.
  • the load pressure of a swing motor becomes high at the beginning of the swing operation and exceeds the limit pressure of a relief valve provided for circuit protection, because a swing body is an inertial body.
  • the boom load pressure which represents a boom holding pressure is lower than the swing load pressure.
  • the swing load pressure exceeds the limit pressure of a relief valve at the beginning of the swing operation as mentioned above.
  • construction machines such as hydraulic excavators desirably have characteristics of the flow control valve which are not determined uniquely for specific pressure compensating and/or flow distributing function, but can be modified to flexibly provide various functions depending on the types of working members and the working modes thereof driven by respective hydraulic actuators.
  • each flow control valve connected to a main circuit between a hydraulic pump and a hydraulic actuator is constituted by a combination of a main valve of the seat valve type, and a pilot valve connected in a pilot circuit between a back pressure chamber of the main valve and an outlet port.
  • An auxiliary valve is also disposed in the pilot circuit, and the inlet and outlet pressures of the pilot valve are introduced to opposite operating parts of the auxiliary valve, respectively, for thereby providing a pressure compensating function.
  • the above patent further discloses a modification in which the self-load pressure is used to affect operation of the single hydraulic actuator for correction of the pressure compensating function.
  • auxiliary valve merely enables to perform a pressure compensating function in connection with the operation of the single hydraulic actuator, or modify the pressure compensating function by introducing an effect of the self-load pressure of the single hydraulic actuator.
  • this known system has no relation with the technique of modifying various functions in the combined operation of a plurality of hydraulic actuators.
  • each of said first and second flow control valve means comprises; a main valve of the seat valve type having a valve body for controlling communication between an inlet port and an outlet port both connected to said main circuit, a variable restrictor capable of changing an opening degree thereof in response to displacements of said valve body, and a back pressure chamber communicating with said inlet port through said variable restrictor and producing a control pressure to urge said valve body in the valve-closing direction; and a pilot circuit connected between said back pressure chamber and said outlet port of said main valve; said first valve means is connected in said pilot circuit as a pilot valve for controlling a pilot flow passing through said pilot circuit, and said second valve means is connected in said pilot circuit as auxiliary valve means for controlling a differential pressure between the inlet pressure and the outlet pressure of said pilot valve; and wherein said control means controls said auxiliary valve means for each of said first and second flow control valve means such that the differential pressure between the inlet pressure and the outlet pressure of said pilot valve has a relationship as
  • the equation has the following meaning.
  • the first term Ps - Pl max at the right side is common to all the flow control valves and hence governs a flow distributing function in the combined operation
  • the second term Pl max - Pl is changed depending on the maximum load pressure among other actuators and hence governs a harmonizing function in the combined operation
  • the third term ⁇ Pl is changed depending on the self-load pressure and hence governs a self-pressure compensating function. Actuation or non-actuation and the degree of these three functions are determined depending on respective values of the constants ⁇ , ⁇ , ⁇ .
  • the flow distributing function represented by the first term is an essential basic function for the combined operation.
  • the constant ⁇ is set to a predetermined positive value irrespective of the types of associated working members.
  • the harmonizing function and the self-pressure compensating function respectively represented by the second and third terms are additional functions effected depending on the types of associated working members and the working modes thereof. Therefore, the constants ⁇ , ⁇ are each set to a predetermined value including zero.
  • the auxiliary valves are not installed in the main circuits but in the pilot circuits, and the main valves installed in the main circuits are constituted in the form of seat valves.
  • the first constant ⁇ preferably meets a relationship of ⁇ ⁇ K, assuming that K is a ratio of the pressure receiving area of the valve body of the main valve, which undergoes the delivery pressure of the hydraulic pump through the inlet port, to the pressure receiving area of the valve body of the main valve, which undergoes the control pressure of the back pressure chamber.
  • K is a ratio of the pressure receiving area of the valve body of the main valve, which undergoes the delivery pressure of the hydraulic pump through the inlet port, to the pressure receiving area of the valve body of the main valve, which undergoes the control pressure of the back pressure chamber.
  • first and second flow control valves have their differential pressures given by the first term at the right side of the above equation substantially equal to each other, so that the flow rate can accurately be distributed in proportion to the operated amounts of the operation means (i.e., opening degrees of the pilot valves) in the fluid distributing function.
  • the first constant ⁇ has the meaning of a proportional gain of the pilot flow rate with respect to the operated amount of the operation means (i.e., opening degree of the pilot valve), namely a proportional gain of the flow rate passing through the main valve with respect to that operated amount.
  • the first constant ⁇ is set to any desired positive value corresponding to the proportional gain.
  • ⁇ K is set, the maximum proportional gain can be provided while attaining the fluid distributing function to distribute the flow rate in proportion to the operated amounts of the operation means.
  • the second constant ⁇ is set to any desired value taking into account harmonization in the combined operation of the associated hydraulic actuator and one or more other hydraulic actuators.
  • ⁇ is set equal to zero.
  • the third constant ⁇ is set to any desired value taking into account operating characteristics of the associated hydraulic actuator.
  • ⁇ is also set equal to zero.
  • the control means may have a plurality of hydraulic control chambers provided in each of the auxiliary valve for the first and second flow control valve means, and line means for directly or indirectly introducing the delivery pressure of the hydraulic pump, the maximum load pressure, and the inlet pressure and the outlet pressure of the pilot valve to the plurality of hydraulic control chambers.
  • the respective pressure receiving areas of the plurality of hydraulic control chambers are set such that the first, second and third constants ⁇ , ⁇ , ⁇ take the respective predetermined values.
  • the auxiliary valve is disposed between the back pressure chamber of the main valve and the pilot valve
  • the plurality of hydraulic control chambers comprise a first hydraulic control chamber for urging the auxiliary valve in the valve-opening direction, and second, third and fourth hydraulic control chambers for urging the auxiliary valve in the valve-closing direction
  • the line means comprises a first line for introducing the control pressure in the back pressure of the main valve to the first hydraulic chamber, a second line for introducing the inlet pressure of the pilot valve to the second hydraulic control chamber, a third line for introducing the maximum load pressure to the third hydraulic control chamber, and a fourth line for introducing the delivery pressure of the hydraulic pump to he fourth hydraulic control chamber.
  • the first and second flow control valves can each be constituted by incorporating the main valve and the auxiliary valve into an integral structure. This provides a compact and rational valve structure.
  • control means may comprise electromagnetic operating parts provided in each of the auxiliary valve means for the first and second flow control valves, pressure detector means for directly or indirectly detecting the delivery pressure of the hydraulic pump, the maximum load pressure, and the inlet pressure and the outlet pressure of the pilot valve, and processing means for calculating a differential pressure between the inlet pressure and the outlet pressure of the pilot valve based on detected signals from the pressure detector means, and then outputting a calculated differential pressure signal to the electromagnetic operating parts of the auxiliary valve means.
  • the first, second and third constants ⁇ , ⁇ , ⁇ are preset as the respective predetermined values in the processing means.
  • the pump control means can be a pump regulator of load sensing type for holding the delivery pressure of the hydraulic pump higher a predetermined value than the maximum load pressure between the first and second hydraulic actuators.
  • Fig.1 is a schematic view showing an overall arrangement of a hydraulic drive system according to one embodiment of the present invention.
  • Fig. 2 is a sectional view showing the structure of a flow control valve of the hydraulic drive system.
  • Fig. 3 is a side view of a hydraulic excavator to which the hydraulic drive system of the present invention is to be applied.
  • Fig. 4 is a plan view of the hydraulic excavator.
  • Fig. 5 is a characteristic graph showing a setting example of the constant ⁇ for a pressure compensating valve included in one flow control valve of the hydraulic drive system.
  • Figs. 6(A) through 6(D) are characteristic graphs each showing a setting example of the constant ⁇ for a pressure compensating valve included in one flow control valve of the hydraulic drive system.
  • Figs. 7(A) through 7(C) are characteristic graphs each showing a setting example of the constant ⁇ for a pressure compensating valve included in one flow control valve of the hydraulic drive system.
  • Fig. 8 is a schematic view showing an overall arrangement of a hydraulic drive system according to another embodiment of the present invention.
  • Fig. 9 is a sectional view showing the structure of a flow control valve of the hydraulic drive system of Fig. 8.
  • Fig. 10 is a sectional view showing a modification of the flow control valve of Fig. 9.
  • Fig. 11 is a sectional view showing another modification of the flow control valve of Fig. 9.
  • Fig. 12 is a schematic view showing an overall arrangement of a hydraulic drive system according to still another embodiment of the present invention.
  • Fig. 13 is a sectional view showing the structure of a flow control valve of the hydraulic drive system of Fig. 12.
  • Figs. 14 through 20 are schematic views showing respective flow control valves of hydraulic drive systems according to still other embodiments of the present invention.
  • Fig. 21 is a schematic view showing an overall arrangement of a hydraulic drive system according to still another embodiment of the present invention.
  • Fig. 22 is a schematic view showing an arrangement of a control unit of the hydraulic drive system of Fig. 21.
  • Fig. 23 is a flowchart showing the procedure of generating a control signal in the control unit.
  • Fig. 24 is a sectional view showing an embodiment in which a main valve and a pressure compensating valve of the flow control valve for use in the hydraulic drive system of the present invention are incorporated into an integral structure.
  • Fig. 25 is a circuit diagram showing an embodiment of a pump regulator of load sensing type where a fixed displacement pump is used in the hydraulic drive system of the present invention.
  • Fig. 26 is a circuit diagram showing an embodiment of pump control means of not load sensing type which is used in the hydraulic drive system of the present invention.
  • a hydraulic drive system comprises a variable delivery hydraulic pump 1 of swash plate type, for example, a plurality of hydraulic actuators 6, 7 connected to the hydraulic pump 1 through main lines 2, 3 and 4, 5 which serve as main circuits, respectively, and driven by hydraulic fluid delivered from the hydraulic pump 1, and flow control valves 8, 9 connected to the main lines 2, 3 and 4, 5 between the hydraulic pump 1 and the hydraulic actuators 6, 7, respectively.
  • the hydraulic pump 1 is associated with a pump regulator 10 of load sensing type which serves to hold a delivery pressure of the hydraulic pump 1 higher a predetermined value than a maximum load pressure among the plurality of hydraulic actuators 6, 7.
  • the flow control valve 8 comprises a main valve 11 connected to the main lines 2, 3 between the hydraulic pump 1 and the hydraulic actuator 6, pilot lines 12, 13, 14 jointly constituting a pilot circuit for the main value 11, a pilot valve 15 connected to the pilot lines 13, 14, and a pressure compensating valve 16, as an auxiliary valve, connected to the pilot lines 12, 13 in series with the pilot valve 15.
  • the main valve 11 comprises a valve housing 19 having an inlet port 17 and an output port 18 connected to the main lines 2, 3, respectively, and a valve body 21 disposed in the valve housing 19 and engageable with a valve seat 20, to thereby control communication between the inlet port 17 and the outlet port 18 in response to displacements (i.e., opening degrees) of the valve body 21 with respect to the valve seat 20.
  • the valve body 21 has formed in its outer circumference a plurality of axial slits 22, the slits 22 being cooperable with the inner wall of the valve housing 19 to constitute a variable restrictor 23 which has a variable opening degree in response to displacements of the valve body 21.
  • a back pressure chamber 24 communicating with the inlet port 17 through the variable restrictor 23 and producing a control pressure Pc.
  • the upper annular end surface (as viewed on the drawing sheet) of the valve body 21 facing the inlet port 17 defines an annular pressure receiving area As which receives the delivery pressure Ps of the hydraulic pump 1
  • the bottom wall surface of the valve body 21 facing the output port 18 defines a pressure receiving area Al which receives a load pressure Pl of the hydraulic actuator 6,
  • the top end surface of the valve body 21 facing the back pressure chamber 24 defines a pressure receiving area Ac which receives a control pressure Pc.
  • Ac As + Al .
  • the pilot line 12 is connected to the back pressure chamber 24 of the main valve 11, and the pilot line 14 is connected to the outlet port 18 of the main valve.
  • the pilot valve 15 comprises a control lever 30, and a valve body 33 of needle type driven by the control lever 30 for controlling communication between an inlet port 31 connected to the pilot line 13 and an outlet port 32 connected to the pilot line 14.
  • the pressure compensating valve 16 comprises a valve body 42 of spool type for controlling communication between an inlet port 40 connected to the pilot line 12 and an outlet port 41 connected to the pilot line 13, first and second hydraulic control chambers 43, 44 for urging the valve body 42 in the valve-opening direction, and third and fourth hydraulic chambers 45, 46 positioned in opposite relation to the first and second hydraulic control chambers 43, 44 for urging the valve body 42 in the valve-closing direction.
  • the first hydraulic control chamber 43 is connected to the main line 2 through a pilot line 47
  • the second hydraulic is control chamber 44 is connected to the pilot line 14, i.e., the outlet side of the pilot valve 15, through a pilot line 48
  • the third hydraulic control chamber 45 is connected to a maximum load pressure line 50 (described later on) through a pilot line 49
  • the fourth hydraulic control chamber 46 is connected to the pilot line 13, i.e., the inlet side of the pilot valve 15, through a pilot line 51.
  • the pilot line 51 is formed as an inner passage of the valve body 42.
  • the delivery pressure Ps of the hydraulic pump 1 is introduced to the first hydraulic control chamber 43
  • the outlet pressure Pl of the pilot valve 15 is introduced to the second hydraulic chamber 44
  • the inlet pressure Pz of the pilot valve 15 is introduced to the fourth hydraulic control chamber 46
  • the load pressure of either hydraulic actuator 6 or 7 on the higher pressure side, i.e. the maximum load pressure Pl is introduced to the third hydraulic control chamber 45.
  • the end surface of the valve body 42 facing the first hydraulics control chamber 43 defines a pressure receiving area as which receives the delivery pressure Ps of the hydraulic pump 1
  • the annular end surface thereof facing the second hydraulic control chamber 44 defines a pressure receiving area al which receives the outlet pressure Pl of the pilot valve 15
  • the end surface thereof facing the third hydraulic control chamber 45 defines a pressure receiving area am which receives the load pressure of either hydraulic actuator 6 or 7 on the higher pressure side, i.e., the maximum load pressure Pl max
  • the annular end surface thereof facing the fourth hydraulic control chamber 46 defines a pressure receiving area az which receives the inlet pressure Pz of the pilot valve 15.
  • setting of the first, second and third constants ⁇ , ⁇ , ⁇ to the respective predetermined values is made by properly selecting the pressure receiving areas as, al, am, az of the first through fourth hydraulic control chambers 43-46.
  • the pressure receiving areas as, al, am, az of the first through fourth hydraulic control chambers 43-46 are so set as to obtain the respective predetermined values of the first, second and third constants ⁇ , ⁇ , ⁇ .
  • the pressure receiving areas as, al, am, az of the first through fourth hydraulic control chambers 43-46 are set such that the valve body 42 is held at its open position so long as the main valve 11 and the pilot valve 15 are being closed.
  • the combination of the main valve 11 of seat value type and the pilot valve 15 is known from U.S.P. No. 4,535,809.
  • a pilot flow is formed in the pilot circuit 12-14 in response to the opening degree of the pilot valve 15.
  • the valve body 21 of the main valve is opened to the opening degree proportional to the pilot flow rate, so that the flow rate corresponding to the operated amount of the control lever 30 (i.e., opening degree of the pilot valve 15) is passed from the inlet port 17 to the outlet port 18 through the main valve 11.
  • the flow control valve 9 is constructed similarly to the flow control valve 8, and comprises a main valve 70 of seat valve type, pilot lines 71, 72, 73 jointly constituting a pilot circuit, a pilot valve 74, and a pressure compensating valve 75.
  • the pilot lines 14, 73 of the flow control valves 8, 9 are connected to the maximum load pressure line 50 through load pressure introducing lines 54, 55 which have check values 52, 53 therein, respectively,
  • the load pressure of either hydraulic actuator 6 or 7 on the higher pressure side is introduced as a maximum load pressure to the maximum load pressure line 50.
  • the maximum load pressure line 50 is connected to a tank 57 through a restrictor 56.
  • check values 58, 59 for preventing hydraulic fluid from flowing toward the main valves 11, 70 from the hydraulic actuators 6, 7 are connected to the main lines 3, 5 downstream of the main valves 11, 70 of the flow control valves 8, 9, respectively.
  • the pump regulator 10 comprises an auxiliary pump 60, a swash plate tilting device 61 of hydraulic cylinder type driven by hydraulic fluid delivered from the auxiliary pump 60, and a control valve 62 connected between the tank 57 as well as the auxiliary pump 60 and the swash plate tilting device 61.
  • the control valve 62 has first and second pilot chambers 63, 64 at its opposite ends, and a pressure setting spring 65 disposed at the end near the second pilot chamber 64.
  • the first and second pilot chambers 63, 64 are connected to the main line 2 and the maximum load pressure line 50 through pilot lines 66, 67, respectively.
  • control valve 62 receives the delivery pressure of the hydraulic pump 1 and the maximum load pressure plus a resilient force of the spring 65 in opposite directions, so that supply and discharge of hydraulic fluid with respect to the swash plate tilting device 61 are controlled in response to changes in the maximum load pressure. Consequently, the delivery pressure of the hydraulic pump 1 is held at a higher pressure than the maximum load pressure by a preset pressure corresponding to the resilient strength of the spring 65.
  • the equation (1) will be taken into consideration below.
  • the left side of the equation (1) relates to a differential pressure ⁇ Pz between the inlet pressure Pz and the outlet pressure Pl of the pilot valve 15.
  • the first term in the right side of the equation (1) relates to a differential pressure between the delivery pressure Ps of the hydraulic pump 1 and the maximum load pressure Pl max, with ⁇ being a proportional constant.
  • the second term relates to a differential pressure between the maximum load pressure Pl max and the load pressure of either the hydraulic actuator 6 or 7, i.e., self-load pressure Pl, with ⁇ being a proportional constant.
  • the third term is determined by the self-load pressure Pl with ⁇ being a proportional constant.
  • the equation (1) means that each of the pressure compensating valves 16, 75 can control the differential pressure ⁇ Pz between the inlet pressure Pz and the outlet pressure Pl of the pilot valve 15 based on the four pressures Ps, Pl max, Pl, Pz; that at this time, the differential pressure ⁇ Pz can be controlled in proportion to such three elements as the differential pressure Ps - Pl max between the delivery pressure Ps of the hydraulic pump 1 and the maximum load pressure Pl max, the differential pressure Pl max - Pl between the maximum load pressure Pl max and the self-load pressure Pl, and the self-load pressure Pl, respectively; and that the respective degrees of proportion to those three elements Ps - Pl max, Pl max - Pl and Pl can optionally be set by selecting respective values of the proportional constants ⁇ , ⁇ , ⁇ .
  • differential pressure Ps - Pl max in the first term in the right side of the equation (1) remains constant in this embodiment using the pump regulator 10 of load sensing type, so long as the pump regulator 10 is working effectively. That differential pressure is common to both the pressure compensating valves 16, 75.
  • controlling the differential pressure ⁇ Pz across the pilot valve 15, 74 in proportion to the differential pressure Ps - Pl max means that the differential pressure ⁇ Pz is controlled at constant in the operating condition where the pump regulator 10 is working effectively.
  • the opening degree of the pilot valve 15, 74 to be constant, it also means that the main flow rate passing through the main valve 11, 70 is controlled at constant irrespective of fluctuations in the inlet pressure Ps or the outlet pressure Pl of the main valve. In short, the pressure compensating function is performed.
  • the flow rates passing through the main valves 11, 70 are distributed proportionally in response to the respective operated amounts of the control levers 30 (i.e., opening degrees of the pilot valves 15, 74), so that the delivery flow rate of the hydraulic pump 1 is reliably supplied to the hydraulic actuator on the higher pressure side as well.
  • the flow distributing function can be attained.
  • controlling the differential pressure ⁇ Pz across the pilot valve 15, 74 in proportion to the differential pressure Pl max - Pl means that when the load pressure Pl max of the other hydraulic actuator is larger than the self-load pressure Pl, the differential pressure ⁇ Pz across the pilot valve 15 or 74 is changed depending on the maximum load pressure Pl max of the other hydraulic actuator. Assuming the opening degree of the pilot valve 15 or 74 to be constant, it also means that the main rate passing through the main valve 11, 70 is changed depending on he maximum load pressure Pl max.
  • While preferred flow control is generally effected by the flow control valves free of any effects from other hydraulic actuators, it may be preferable in hydraulic construction machines such as hydraulic excavators to vary the respective flow rates under the effects of load pressures of other hydraulic actuators depending on the working modes.
  • the second term in the right side of the equation (1) represent a harmonizing function with which the respective flow rates can be changed for harmonization with other hydraulic actuators.
  • controlling the differential pressure ⁇ Pz across the pilot valve 15, 74 in proportion to the self-load pressure Pl means that the differential pressure ⁇ Pz across the pilot valve 15 or 74 is changed in response to changes in the self-load pressure Pl .
  • the opening degree of the pilot valve 15 or 74 to be constant, it also means that the main flow rate passing through the main valve 11, 70 is changed depending on the self-load pressure Pl. This provides a self-pressure compensating function with which the flow rate can be varied in response to changes in the self-load pressure.
  • the first term in the right side of the equation (1) governs the pressure compensating and flow distributing function
  • the second term governs the harmonizing function in combination with other hydraulic actuators
  • the third term governs the self-pressure compensating function. Actuation or non-actuation and the degree of each of those three functions can optionally be set by selecting the proportional constants ⁇ , ⁇ , ⁇ .
  • the pressure compensating and flow distributing function in relation to the first term is an essential function for hydraulic construction machines such as hydraulic excavators, and is preferably held constant at all times irrespective of the types and working forms of hydraulic actuators employed. Therefore, the proportional constant ⁇ is set to an arbitrary positive value.
  • the proportional constant ⁇ for the differential pressure Pl max - Pl of the first term means a proportional gain of the pilot flow rate with respect to the operated amount of the control lever 30 for the pilot valve 15, 74 (opening degree of the pilot valve), i.e., a proportional gain of the main flow rate passing through the main valve 11, 70 with respect to that operated amount. Therefore, the proportional constant ⁇ is determined corresponding to such proportional gain.
  • the maximum differential pressure obtainable with the pilot valve 15, 74 is K (Ps - Pl) .
  • the proportional constant ⁇ should be set to meet ⁇ ⁇ K.
  • ⁇ is set to meet ⁇ ⁇ K
  • the differential pressure across the pilot valve on the side of lower load pressure is limited to K(Ps - Pl max) also in the combined operation. Even when the combined operation is switched to the sole operation, no variation occurs in the differential pressure, thereby ensuring the stable work operation. Therefore, also from this viewpoint, ⁇ is preferably set to meet ⁇ ⁇ K.
  • the harmonizing function relating to the second term has different degrees of necessity depending on the types of working members and the working modes driven by the hydraulic actuators 6, 7 It is preferable for some working members and modes to be totally unaffected from the load pressure of the other hydraulic actuator. Therefore, the proportional constant ⁇ is set to an arbitrary value inclusive of zero based on harmonization in the combined operation of the relevant hydraulic actuator with the other hydraulic actuator.
  • the self-pressure compensating function relating to the third team has different degrees of necessity depending on the types of working members driven by the hydraulic actuators 6, 7. It is also preferable for some working members to be totally unaffected from the self-load pressure. Therefore, the proportional constant ⁇ is set to an arbitrary value inclusive of zero depending on the types of working members driven by the relevant hydraulic actuator.
  • the proportional constants ⁇ , ⁇ , ⁇ are expressed using the pressure receiving areas as, al, am, az of the first through fourth hydraulic control chambers 43-46 of the pressure compensating valve 16, 75. Accordingly, if the proportional constants ⁇ , ⁇ , ⁇ are once determined, the pressure receiving areas as, al, am, az are so set as to obtain those determined values of the proportional constants ⁇ , ⁇ , ⁇ .
  • a hydraulic excavator generally comprises a pair of track bodies 80, a swing body 81 swingably installed on the track bodies 80, and a front attachment 82 mounted onto the swing body 81 rotatably in a vertical plane.
  • the front attachment 82 comprises a boom 83, an arm 84 and a bucket 85.
  • the track bodies 80, swing body 81, boom 83, arm 84 and bucket 85 are driven by a plurality of track motors 86, swing motor 87, boom cylinder 88, arm cylinder 89 and bucket cylinder 90, respectively.
  • the swing motor 87, boom cylinder 88, arm cylinder 89 and bucket cylinder 90 correspond each to one or more hydraulic actuators 6, 7 shown in Fig. 1.
  • the proportional constants ⁇ of the first term commonly affecting to all flow control valves of the swing motor 87, boom cylinder 88, arm cylinder 89 and bucket cylinder 90 are set to the same arbitrary positive value taking into account the above-mentioned proportional gain, as shown in Fig. 5 by way of example.
  • the proportional constant ⁇ is set to a relatively small negative value as shown in Fig.
  • the proportional constant ⁇ is set to a relatively small positive value as shown in Fig. 7(C).
  • the proportional constants ⁇ , ⁇ are all set to zero as shown in Figs. 6(A) and 7(B).
  • the pilot valves 15, 74 are closed and, hence, no pilot flow rates pass through the pilot circuits 12-14, 71-73. Therefore, hydraulic fluid will not flow through the respective variable restrictors 23 of the main valves 11, 70, so the control pressure Pc of the back pressure chamber 24 is equal to the pressure Ps at the inlet port 17 (i.e., delivery pressure of the hydraulic pump 1). Further, due to the above-mentioned action of the pump regulator 10 of load sensing type, the delivery pressure Ps of the hydraulic pump 1 is held at a pressure level higher than the maximum load pressure Pl max between the hydraulic actuators 6, 7 by an amount of pressure corresponding to a preset value of the spring 65.
  • the pilot valve 15 is opened in response to the operated amount thereof to produce a pilot flow in the pilot circuit 12-14, so the pilot flow rate passes corresponding to the opening degree of the pilot valve 15.
  • this causes the valve body 21 of the main valve to be opened to an opening degree proportional to the pilot flow rate under the action of the variable restrictor 23 and the back pressure chamber 24.
  • the flow rate corresponding to the operated amount of the control lever 30 i.e., opening degree of the pilot valve 15
  • the pump regulator 10 of load sensing type functions to increase the delivery pressure of the hydraulic pump 1, so that the differential pressure between the pressure at the inlet port 17 (i.e., delivery pressure of the hydraulic pump 1) and the pressure at the outlet port 18 (i.e., load pressure of the hydraulic actuator 6; maximum load pressure) is held constant. Therefore, the certain flow rate corresponding the operated amount of the control lever 30 still continues to pass through the main valve 11.
  • the proportional constant ⁇ for the flow control valve associated with the swing motor 87 is set to a negative value near zero as shown in Fig. 7(A). More specifically, when driving the swing body 81, the load pressure is increased beyond the limit pressure of a relief valve provided to protect the circuit since the swing body is of an inertial body. This results in waste of energy.
  • the proportional constant ⁇ is set to a negative value, the differential pressure Pz - Pl is controlled to be reduced with increasing the load pressure of the swing body, thereby reducing the flow rate passing through the flow control valve. This makes smaller the amount of flow rate dissipated away as a surplus flow rate from the relief valve even if the load pressure is raised up, and hence energy is less wasted.
  • the proportional constant ⁇ is set to a small positive value as shown in Fig. 7(C). Accordingly, as the self-load pressure is raised up during the excavation, the differential pressure Pz - Pl is increased to enlarge the flow rate passing through the flow control valve. Thus, the excavation speed of bucket is increased. This enables to give the excavation with powerful feeling and improve operability.
  • the pressure compensating and flow distributing function is carried out by previously setting the pressure receiving areas as, al, am, az of each of the pressure compensating valves 16, 17 such that the proportional constant ⁇ for the first term in the right side of the equation (1) takes an arbitrary positive value as shown in Fig. 5.
  • ⁇ ⁇ K also makes it possible to supply the flow rates to the respective hydraulic actuators accurately in proportion to the operated amounts of the corresponding control levers.
  • the pressure receiving areas as, al, am, az of the pressure compensating valves 16 are selected such that the proportional constants ⁇ , ⁇ in the above equation (1) become zero, the path along which each working member moves can accurately be controlled corresponding to the operated amount of the control lever.
  • the auxiliary valves are installed in not the main circuits but the pilot circuits. Therefore, the fluid leakage is very small even when the hydraulic circuit is highly pressurized, and appreciable pressure loss will not occur if a large flow rate is passed through the main circuit.
  • the harmonizing function and/or the self-load pressure compensating function based on the above pressure compensating and flow distributing function are performed so as to change the main flow rates passing through the main valves 11, 70 depending on the maximum load pressure Pl max among other hydraulic actuators and/or the self-load pressure Pl.
  • the load pressure of the swing motor becomes higher at the initial stage of swing operation since the swing body 81 is of an inertial body.
  • the load pressure is reduced.
  • the load pressure of the boom cylinder is given by a boom holding pressure, it is lower than the load pressure of the swing motor at the initial stage of swing operation.
  • the swing and boom-up operations are actuated in digging work effected by an excavator of backhoe type, for example, it is preferable that even if an operator concurrently operates both the swing and boom-up control levers up to their full strokes for simpler manual operation, the boom-up and swing speeds are automatically adjusted such that the boom-up speed is increased faster than the swing speed at the initial stage and, after the boom has been raised up to some extent, the swing speed is increased gradually.
  • the flow control valve associated with the boom operates in such a manner that during the time the load pressure of the swing motor is high and the differential pressure Pl max - Pl is large at the initial stage of swing operation, the differential pressure ⁇ Pz across the pilot valve is also large to increase the flow rate supplied to the boom cylinder, and thereafter ⁇ Pz is reduced gradually as the differential pressure Pl max - Pl is lowered.
  • the boom-up and swing speeds can be adjusted automatically and the operator can make the manual operation more easily.
  • the proportional constant ⁇ is set to a relatively small positive value as shown in Fig. 6(C).
  • the opening degree of the main valve for the flow control valve associated with the arm is increased in response to rise-up of the differential pressure Pl max - Pl, and hence the restriction degree of hydraulic fluid becomes smaller. This enables to less degrade both fuel economy and heat balance.
  • the proportional constant ⁇ is set to a relatively small negative value as shown in Fig. 6(D).
  • the proportional constant ⁇ is set to a relatively small negative value as shown in Fig. 6(D).
  • the hydraulic drive system of this embodiment can provide the flow distributing function, or the harmonizing function and/or the self-pressure compensating function based on the flow distributing function, and can modify the characteristics of the flow control valves depending on the types of working members of hydraulic construction machines and the working modes thereof, by properly selecting the respective pressure receiving areas of each of the pressure compensating valves and setting the proportional constant ⁇ , ⁇ , ⁇ to their predetermined values.
  • each pressure compensating valve serving as an auxiliary valve is disposed in not the main circuit but the pilot circuit, and each main valve disposed in the main circuit is constituted by a seat valve. Therefore, fluid leakage is very small, which makes the hydraulic circuit more suitable for higher pressurization.
  • the auxiliary valve is disposed in the pilot circuit, appreciable pressure loss will not occur at the auxiliary valve even if a large flow rate is passed through the main circuit. This is also economical.
  • the delivery pressure Ps of the hydraulic pump 1, the maximum load pressure Pl max, and the inlet and outlet pressures Pz, Pl of the pilot valves 15, 74 are directly employed for controlling the pressure compensating valves 16, 75.
  • these four pressures are related to each other via the control pressure of the back pressure chamber 24, so it is also possible to control the pressure compensating valves and provide the above-mentioned characteristics to the respective pressure compensating valves without direct use of all the four pressures.
  • Figs. 8 and 9 shows another embodiment in which the four pressures are not directly employed for controlling the pressure compensating valves from the above standpoint.
  • Fig. 8 and 9 shows another embodiment in which the four pressures are not directly employed for controlling the pressure compensating valves from the above standpoint.
  • FIG. 1 shows only the flow control valves 8, 9 disposed in the meter-in (inlet side) circuit as used when the hydraulic actuators 6, 7 are actuated to be extended or rotated in one direction, the flow control valves 8, 9 function each as a part of a directional control valve in a practical circuit.
  • Fig. 8 shows an overall arrangement of the directional control valve.
  • directional control valves 100, 101 for controlling actuation of hydraulic cylinders 6, 7 are disposed between a hydraulic pump 1 and the hydraulic cylinders 6, 7, respectively, the directional control valve 100 comprising four flow control valves 102, 103, 104, 105 of seat valve type.
  • the first flow control valve 102 is connected to a meter-in (inlet side) circuit 106 as used when the hydraulic cylinder 6 is actuated to be extended, and it corresponds to the flow control valve 8 in the embodiment shown in Fig. 1.
  • the second flow control valve 103 is connected to a meter-in circuit 107 as used when the hydraulic cylinder 6 is actuated to be contracted
  • the third flow control valve 104 is connected to a meter-out (outlet side) circuit 108, between the hydraulic cylinder 6 and the second flow control valve 103, as used when the hydraulic cylinder 6 is actuated to be extended
  • the fourth flow control valve 105 is connected to a meter-out circuit 109, between the hydraulic cylinder 6 and the first flow control valve 102, as used when the hydraulic cylinder 6 is actuated to be contracted.
  • a check valve 110 for preventing hydraulic fluid from reversely flowing toward the first flow control valve is connected between the first flow control valve 102 and the fourth flow control valve 105, while another check valve 111 for preventing hydraulic fluid from reversely flowing toward the second flow control valve is connected between the second flow control valve 103 and the third flow control valve 104.
  • the first through fourth flow control valves 102-105 comprise main valves 112, 113, 114, 115 of seat valve type, pilot circuits 116, 117, 118, 119 associated with the corresponding main valves, and pilot valves 120, 121, 122, 123 connected to the corresponding pilot circuits, respectively.
  • the first and second flow control valves 102, 103 further include respective pressure compensating valves 124, 125 connected to the pilot circuits 116, 117 in series with the pilot valves 120, 121.
  • the structure and function of each of the main valves 112-115 are identical to those of the main valve 11, 70 of the embodiment shown in Fig. 1.
  • pilot valves 120-123 when the pilot valves 120-123 are operated, pilot flow rates corresponding to the opening degrees of the pilot valves are produced in the pilot circuits 116-119, respectively.
  • a valve body 21 of each main valve is opened to an opening degree proportional to the pilot flow rate under the action of a variable restrictor 23 and a back pressure chamber 24, so that a flow rate corresponding to the opening degree of each of the pilot valves 120-123 is passed from an inlet port 17 to an output port 18 through the main valve 11.
  • each of the pilot valves 120-123 is basically identical to the pilot valve 15, 17 shown in Fig. 1 except for that the former has a hydraulic control portion 126.
  • the pressure compensating valve 124 comprises a valve body 130 of spool type, a first hydraulic control chamber 131 for urging the valve body 130 in the valve-opening direction, and second, third and fourth hydraulic chambers 132, 133, 134 positioned in opposite relation to the first hydraulic control chamber 131 for urging the valve body 130 in the valve-closing direction.
  • the first hydraulic control chamber 131 is connected to the back pressure chamber 24 of the main valve 112 through a pilot line 135, the second hydraulic control chamber 132 is defined so as to communicate with an outlet port 41 of the pressure compensating valve 124, the third hydraulic control chamber 133 is connected to a maximum load pressure line 50 through a pilot line 136, and the fourth hydraulic control chamber 134 is connected to the main circuit 106 at the inlet port 17 side of the main valve 112 through a pilot line 137.
  • control pressure Pc of the back pressure chamber 24 is introduced to the first hydraulic control chamber 131
  • the inlet pressure Pz of the pilot valve 120 is introduced to the second hydraulic chamber 132
  • the maximum load pressure Pl max is introduced to the third hydraulic control chamber 133
  • the delivery pressure Ps of the hydraulic pump 1 is introduced to the fourth hydraulic control chamber 134.
  • the end surface of the valve body 130 facing the first hydraulic control chamber 131 defines a pressure receiving area ac which receives the control pressure Pc of the back pressure chamber 24
  • the annular end surface of the valve body 130 facing the second hydraulic control chamber 132 defines a pressure receiving area az which receives the inlet pressure Pz of the pilot valve 120
  • the end surface of the valve body 130 facing the third hydraulic control chamber 133 defines a pressure receiving area am which receives the maximum load pressure Pl max
  • the end surface of the valve body 130 facing the fourth hydraulic control chamber 134 defines a pressure receiving area as which receives the delivery pressure Ps of the hydraulic pump 1.
  • pressure receiving areas as, ac, am, az are so set as to obtain the predetermined values of the proportional constants ⁇ , ⁇ , ⁇ as described later on.
  • the pressure receiving areas as, ac, am, az are set to hold the valve body 130 in an open position while the main valve 112 and the pilot valve 120 are being closed.
  • the pressure compensating valve 125 is constructed similarly to the pressure compensating valve 124.
  • the directional control valve 101 associated with the hydraulic cylinder 7 is constructed similarly to the directional control valve 100.
  • the hydraulic pump 1 is associated with a pump regulator 140 of load sensing type for holding the delivery pressure of the hydraulic pump 1 higher a predetermined valve than a maximum load pressure between the plurality of hydraulic actuators 6, 7.
  • the pump regulator 140 comprises a swash plate tilting device 141 of hydraulic cylinder type and a control valve 142.
  • the swash plate tilting device 141 is driven based on an area difference between a rod side cylinder chamber and a head side cylinder chamber, depending on a position of the control valve 142 for controlling the delivery flow rate of the hydraulic pump 1.
  • the control valve 142 is driven in a like manner to the control valve 62 shown in Fig. 1.
  • control valve 142 undergoes the delivery pressure of the hydraulic pump 1 plus the maximum load pressure and a preset resilient force of a spring 65, which act thereon in opposite directions to control the swash plate tilting device 141 in response to changes in the maximum load pressure, thereby holding the delivery pressure of the hydraulic pump 1 higher than the maximum load pressure by a pressure value corresponding to the resilient strength of the spring 65.
  • the differential pressure ⁇ Pz across the pilot valve 120 can be controlled in proportion to three factors; the differential pressure Ps - Pl max between the delivery pressure Ps of the hydraulic pump 1 and the maximum load pressure Pl max, the differential pressure Pl max - Pl between the maximum load pressure Pl max and the self-load pressure Pl, and the self-load pressure Pl, respectively, thereby enabling to attain the pressure compensating and flow distributing function (first term in the right side), or the harmonizing function (second term in the right side) and/or the self-pressure compensating function (third term in the right side) in the combined operation based on the pressure compensating and flow distributing function, as mentioned above.
  • this embodiment introduces the control pressure Pc, the inlet pressure Pz of the pilot valve 120, the maximum load pressure Pl max and the delivery pressure Ps of the hydraulic pump 1 rather than directly using the inlet and outlet pressures Pz, Pl of the pilot valve 120, the delivery pressure Ps of the hydraulic pump 1 and the maximum load pressure Pl max, in order to provide the same effect as attained using the latter four pressures Pz, Pl, Ps, Pl max.
  • Fig. 10 shows a modification in which the hydraulic control chambers of the pressure compensating valve shown in Fig. 9 is changed in their layout. More specifically, in a pressure compensating valve 150 of this modified embodiment, a first hydraulic control chamber 151 receiving the control pressure Pc of the back pressure chamber 24 is located adjacent to the back pressure chamber 24, the aforesaid pilot line 135 is dispensed with, and three hydraulic control chambers located in opposite relation to the first hydraulic control chamber 151 are arranged in the order of a hydraulic control chamber 152 receiving the inlet pressure Pz of the pilot valve 120, a hydraulic control chamber 153 receiving the delivery pressure Ps of the hydraulic pump 1, and a hydraulic control chamber 154 receiving the maximum load pressure Pl max. With the hydraulic control chambers thus arranged, the above equation (4) is also established, and hence there can be obtained the same effect as with the embodiment shown in Fig. 9.
  • Fig. 11 shows a modified structure of the main valve of seat type.
  • a main valve 160 of seat type includes a valve body 162 having a through hole 161 capable of communicating the inlet port 17 with the back pressure chamber 24, in place of the valve body having the slits 22 as a variable restrictor which was used in the foregoing embodiment.
  • the through hole 161 serves as a variable restrictor adapted to change the restricted amount of hydraulic fluid in response to movement of the valve body 162.
  • the axial direction of the inlet port 17 is normal to the direction of movement of the valve body 21 and the axial direction of the outlet port 18 is aligned with the direction of movement of the valve body 21
  • this modified embodiment is arranged such that the axial direction of the inlet port 17 is aligned with the direction of movement of the valve body 162 and the axial direction of the outlet 18 is normal to the direction of movement of the valve body 162.
  • the bottom end surface of the valve body 162 defines the pressure receiving area As which receives the pump delivery pressure Ps.
  • the direction of flow of hydraulic fluid passing from the inlet port 17 to the outlet port 18 is reversed in contrast with the foregoing embodiment.
  • the main valve 160 in this embodiment also functions in a like manner to the main valve 11 of the foregoing embodiment, so that the main flow rate corresponding to the pilot flow rate can be passed under the action of the variable restrictor provided by the through hole 161 and the back pressure chamber 24.
  • the pressure compensating valve 124 is able to function in a like manner to that in the embodiment of Fig. 9 with the same effect.
  • directional control valves are designated at 170, 171 which are identical in their arrangement to those of the embodiment shown in Fig. 8 except for the structure of pressure compensating valves 172, 173.
  • the pressure compensating valve 172 (173) is different from that in the foregoing embodiment in location of the pilot circuit 116 (117). Specifically, the pressure compensating valve 172 (173) is connected to the pilot circuit 116 (117) between the outlet side of the pilot valve 120 (121) and the outlet port 18 of the main valve 102 (103). Another difference is in pressures which are introduced for controlling the pressure compensating valve. More specifically, the pressure compensating valve 172 (173) comprises a valve body 174 of spool type, a first hydraulic control chamber 175 for urging the valve body 174 in the valve-opening direction, and second and third hydraulic control chambers 176, 177 for urging the valve body 174 in the valve-closing direction.
  • the first hydraulic control chamber 175 is defined so as to communicate with an inlet port 178 of the pressure compensating valve the second hydraulic control chamber 176 is connected to the outlet port 18 of the main valve 102 (103) through a pilot line 179, and the third hydraulic control chamber 177 is connected to the maximum load pressure line 50 through a pilot line 180.
  • the outlet pressure Pz of the pilot valve 120 (121) is introduced to the first hydraulic control chamber 175, the outlet pressure (load pressure) Pl of the main valve 102 (103) is introduced to the second hydraulic control chamber 176, and the maximum load pressure Pl is introduced to the third hydraulic control chamber 177.
  • the end surface of the valve body 174 facing the first hydraulic control chamber 175 defines a pressure receiving area az which receives the outlet pressure Ps of the pilot valve
  • the annular end surface of the valve body 174 facing the second hydraulic control chamber 176 defines a pressure receiving area al which receives the outlet pressure Pl of the main valve
  • the end surface of the valve body 174 facing the third hydraulic control chamber 177 defines a pressure receiving area am which receives the maximum load pressure Pl max.
  • These pressure receiving area az, al, am are so set as to obtain the predetermined values of proportional constants ⁇ , ⁇ , ⁇ as described later on.
  • the pressure receiving areas az, al, am are set such that the valve body 174 is held in an open position while the main valve 102 (103) and the pilot valve 120 (121) are being closed.
  • the differential pressure ⁇ Pz across the pilot valve 120 can be controlled in proportion to three factors; the differential pressure Ps - Pl max between the delivery pressure Ps of the hydraulic pump 1 and the maximum load pressure Pl max, the differential pressure Pl max - Pl between the maximum load pressure Pl max and the self-load pressure Pl, and the self-load pressure Pl, respectively, thereby enabling to attain the pressure compensating and flow distributing function (first term in the right side), or the harmonizing function (second term in the right side) and/or the self-pressure compensating function (third term in the right side) in the combined operation based on the pressure compensating and flow distributing function, as mentioned above.
  • the equation (5) can be expressed using four pressure; the inlet and outlet pressures Pc, Pz, the delivery pressure Ps of the hydraulic pump 1, and the maximum load pressure Pl max.
  • this embodiment introduces three pressures i.e., the outlet pressure Pz, the outlet pressure Pl of the main valve, and the maximum load pressure Pl max, rather than directly using the inlet and outlet pressures Pz, Pl of the pilot valve 120, the delivery pressure Ps of the hydraulic pump 1 and the maximum load pressure Pl max, in order to provide the same effect as attained using the latter four pressures Pz, Pl, Ps, Pl max.
  • the present invention is to control each pressure compensating valve based on four pressures, i.e., the inlet and outlet pressures of the pilot valve, the delivery pressure of the hydraulic pump 1, and the maximum load pressure, thereby making it possible to selectively achieve the pressure compensating and flow distributing function, or the harmonizing function and/or the self-pressure compensating function based on the pressure compensating and flow distributing function.
  • These four pressures are related to each other via the control pressure Pc of the back pressure chamber 24, so the pressure compensating valve can also be controlled without direct use of all the four pressures.
  • the pressure compensating valve may be located either upstream or downstream of the pilot valve. Still other modifications in this respect will be explained below. Note that, in the following description, the main valve and the pilot valve are represented by 11, 15, respectively.
  • Fig. 14 shows a modification in which a pressure compensating valve 190 is disposed in the pilot circuit between the back pressure chamber 24 and the pilot valve 15.
  • the control pressure Pc of the back pressure chamber and the outlet pressure Pl of the pilot valve are introduced to the hydraulic control chambers having their pressure receiving areas ac, al and urging the pilot valve in the valve-opening direction, respectively, while the inlet pressure Pz of the pilot valve and the maximum load pressure Pl max are introduced to the hydraulic control chambers having their pressure receiving areas az, am and urging the pilot valve in the valve-closing direction, respectively.
  • Fig. 15 shows a modification in which a pressure compensating valve 191 is disposed between the pilot valve 15 and the outlet port of the main valve 11.
  • the delivery pressure Ps of the hydraulic pump 1 and the outlet pressure Pz of the pilot valve are introduced to the hydraulic control chambers having their pressure receiving areas as, az and urging the pilot valve in the valve-opening direction, respectively, while the inlet pressure Pc of the pilot valve and the maximum load pressure Pl max are introduced to the hydraulic control chambers having their pressure receiving areas ac, am and urging the pilot valve in the valve-closing direction, respectively.
  • Fig. 16 shows a modification in which a pressure compensating valve 192 is disposed between the pilot valve 15 and the outlet port of the main valve 11.
  • the delivery pressure Ps of the hydraulic pump 1 and the outlet pressure Pz of the pilot valve are introduced to the hydraulic control chambers having their pressure receiving areas as, az and urging the pilot valve in the valve-opening direction, respectively, while the maximum load pressure Pl max is introduced to the hydraulic control chamber having its pressure receiving areas am and urging the pilot valve in the valve-closing direction.
  • Fig. 17 shows a modification in which a pressure compensating valve 193 is disposed between the pilot valve 15 and the outlet port of the main valve 11.
  • the delivery pressure Ps of the hydraulic pump 1, the inlet pressure Pc of the pilot valve and the outlet pressure Pz of the pilot valve are introduced to the hydraulic control chambers having their pressure receiving areas as, ac, az and urging the pilot valve in the valve-opening direction, respectively, while the maximum load pressure Pl max is introduced to the hydraulic control chamber having its pressure receiving area am and urging the pilot valve in the valve-closing direction.
  • Fig. 18 shows a modification in which a pressure compensating valve 194 is disposed between the pilot valve 15 and the outlet port of the main valve 11.
  • the outlet pressure Pz of the pilot valve is introduced to the hydraulic control chamber having its pressure receiving area as and urging the pilot valve in the valve-opening direction, while the inlet pressure Pc of the pilot-valve, the outlet pressure Pl of the main valve 11 and the maximum load pressure Pl max are introduced to the hydraulic control chambers having their pressure receiving areas ac, al, am and urging the pilot valve in the valve-closing direction, respectively.
  • Fig. 19 shows a modification in which a pressure compensating valve 195 is disposed between the pilot valve 15 and the outlet port of the main valve 11.
  • the inlet pressure Pc of the pilot valve and the outlet pressure Pz of the pilot valve are introduced to the hydraulic control chambers having their pressure receiving areas ac, as and urging the pilot valve in the valve-opening direction, while the outlet pressure Pl of the main valve 11 and the maximum load pressure Pl max are introduced to the hydraulic control chambers having their pressure receiving areas al, am and urging the pilot valve in the valve-closing direction, respectively.
  • Fig. 20 shows a modification in which a pressure compensating valve 196 is disposed between the pilot valve 15 and the outlet port of the main valve 11.
  • the outlet pressure Pz of the pilot valve, the delivery pressure Ps of the hydraulic pump 1, and the outlet pressure Pl of the main valve 11 are introduced to the hydraulic control chambers having their pressure receiving areas az, as, al and urging the pilot valve in the valve-opening direction, respectively, while the maximum load pressure Pl max is introduced to the hydraulic control chamber having its pressure receiving area am and urging the pilot valve in the valve-closing direction.
  • control means for the pressure compensating valve is constituted by hydraulic means which directly or indirectly introduces the delivery pressure of the hydraulic pump, the maximum load pressure, and the inlet and outlet pressures of the pilot valve to a plurality of hydraulic control chambers, that control means can also be constituted in an electrical manner.
  • Figs. 21 through 23 illustrate one of such embodiments.
  • flow control valves for controlling hydraulic actuators 6, 7 are designated at reference numerals 200, 201, respectively.
  • the flow control valves 200, 201 include pressure compensating valves 202. 203 comprising electromagnetic proportional valves 202, 203 having electromagnetic operating parts 202A, 202B, respectively. Except for this, each of the flow control valves 200, 201 is constructed in the same manner as the flow control valve 8, 9 in the embodiment of Fig. 1.
  • a pressure detector 204 for detecting the delivery pressure Ps of a hydraulic pump 1 is connected to a delivery line of the hydraulic pump 1 in communication with main lines 2, 3, pressure detectors 205, 206 for detecting the inlet pressure Pz of pilot valves 15, 74 are connected to lines 13, 72 of the pilot circuit, respectively, pressure detectors 207, 208 for detecting the outlet pressure Pl of the pilot valves 15, 74 are connected to pilot lines 14, 73, respectively, and a pressure detector 209 for detecting the maximum load pressure Pl max of the hydraulic actuators 6, 7 is connected to a maximum load pressure line 50.
  • the hydraulic pump 1 is associated with an angle gauge 210 for detecting a tilting angle of a swash plate, for example, which is used in a variable displacement mechanism.
  • the delivery flow rate of the hydraulic pump 1 is controlled by a delivery flow control device 212 driven with hydraulic fluid from an auxiliary pump 211.
  • Pressure signals Pz1, Pz2, Pl 1, Pl 2, Pl max from the pressure detectors 204- 209 and a tilting angle signal Qr from the angle gauge 210 are input to a control unit 213 which calculates a control signal Qo for the hydraulic pump 1 and control signals I1o, I2o for the pressure compensating valves 202, 203, and then output these signals to the delivery flow control device 212 and electromagnetic operating parts 202A, 203A of the pressure compensating valves, respectively.
  • the control unit 213 is constituted by a microcomputer and, as shown in Fig. 22, it comprises an A/D converter 214 for converting the above pressure signals and tilting angle signal to digital signals, a central processing unit 215, a memory 216 for storing the program of control procedure, a D/A converter 217 for outputting analog signals, an I/O interface 218, amplifiers 219, 220 connected to the electromagnetic operating parts 202A, 203A of the respective pressure compensating valves, and amplifiers 221, 222 connected to input terminals 212A, 212B of the delivery amount control device 212, respectively.
  • the control unit 213 calculates a delivery flow target value Qo of the hydraulic pump 1 which is effective to hold the pump delivery pressure higher a predetermined value than the maximum load pressure, using the pressure signal Ps from the pressure detector 204 for detecting the delivery pressure of the hydraulic pump 1 and the pressure signal Pl max from the pressure detector 209 for detecting the maximum load pressure between the hydraulic actuators 6, 7 on the basis of the control procedure program stored in the memory 216.
  • the target value signal Qo is output from the amplifiers 221, 222 to the input terminals 212A, 212B of the delivery flow control device 212 through the I/O interface 218.
  • the delivery flow control device 212 controls a tilting angle of the swash plate of the hydraulic pump 1 so that the tilting angle Qr detected by the angle gauge 210 becomes equal to the target value Qo. This holds the pump delivery pressure higher a predetermined value than the maximum load pressure, thereby providing a similar function to the hydraulic pump regulator of load sensing type used in the foregoing embodiments.
  • the control unit 213 also calculates control amounts of the pressure compensating valves 202, 203 based on the pressure signals Ps, Pz1, Pz2, Pl 1, Pl 2, Pl max from the pressure detectors 204 - 209 for controlling the pressure compensating valves.
  • Fig. 23 is a flowchart showing the control procedure of the pressure compensating valves.
  • the microcomputer reads in the pressure signals Ps, Pz1, Pz2, Pl 1, Pl 2, Pl max detected by the pressure detectors 204 -209.
  • step 231 target inlet pressures Pz10, Pz20 of the pilot valves 15, 74 are calculated from the following equations: It is to be noted that these equations are identical to the equation (1) in the first embodiment, the constants ⁇ , ⁇ , ⁇ are set to their predetermined values as shown in Figs. 5 through 7, for example, depending on selection of three functions, i.e., the pressure compensating and flow distributing function, the harmonizing function, and the self-pressure compensating function.
  • I1o G (Pz10 - Pz1)
  • the calculated control signals I1o, I2o are output from the amplifiers 219, 220 to the electromagnetic operating parts 202A, 203A of the pressure compensating valves 202, 203 through the D/A converter 217, respectively.
  • the constants ⁇ , ⁇ , ⁇ are preset as a part of the program.
  • a regulator 240 operable from the exterior may be connected to the control unit 213 as indicated by the imaginary line in Fig. 21, so that the constants ⁇ , ⁇ , ⁇ may variably be set depending on the types of hydraulic construction machines and the working members thereof, etc.
  • Fig. 24 shows an embodiment in which the seat type main valve and the pressure compensating valve of the flow control valve are incorporated into an integral structure.
  • a flow control valve 270 comprises a main valve section 271 and a pressure compensating valve section 272.
  • the main valve section 271 is disposed in a valve housing 275 having an inlet port 273 and an output port 274, and has a valve body 276 of seat valve type for controlling communication between the inlet port 273 and the outlet port 274.
  • the valve body 276 has formed in its circumference a passage 277 which constitutes a variable restrictor, and a back pressure chamber 278 communicating with the inlet port 273 through the variable restrictor 277 is defined at the back of the valve body 276.
  • the pressure compensating valve section 272 has a valve body 280 of spool type disposed in the valve housing 275 for restricting the passage between the back pressure chamber 278 and a pilot outlet port 279.
  • the valve body 280 is engaged with a piston 281 inserted in the valve body 276 of the main valve movably in the axial direction.
  • the pressure compensating valve section 272 also comprises a first hydraulic control chamber 282 in facing relation to the end surface of the valve body 280 opposite to the piston, a second hydraulic control chamber 283 in facing relation to the first annular end surface of the valve body 280, a third hydraulic control chamber 284 in facing relation to the second annular end surface of the valve body 280, and a fourth hydraulic control chamber 285 defined in the valve body 276 of the main valve in facing relation to the end surface of the piston 281.
  • the first hydraulic control chamber 282 is communicated with the back pressure chamber 278 through a passage 286, the second hydraulic control chamber 283 is communicated with a pilot outlet port 279, the third hydraulic control chamber 284 is communicated with a maximum load pressure port 287, and the fourth hydraulic control chamber 285 is communicated with the inlet port 273 of the main valve through a passage 288.
  • the pilot outlet port 279 is connected to a pilot valve 290 through a pilot line 289, and the maximum load pressure port 287 is connected to a maximum load pressure line (not shown).
  • first through fourth hydraulic control chambers 282 - 285 correspond to the first through fourth hydraulic control chambers 131 - 134 of the flow control valve shown in Fig. 9, respectively.
  • the compact and rational valve structure can be obtained by combining the main valve and the pressure compensating valve into an integral structure.
  • the hydraulic drive system was described in combination with the pump regulator of load sensing type, and the pump regulator of load sensing type was described as an implement to control the delivery pressure of the variable displacement hydraulic pump.
  • the hydraulic pump may be of a fixed displacement type.
  • the pump regulator of load sensing type is constructed as shown in Fig. 25. More specifically, in Fig. 25, a pump regulator 380 is associated with a relief valve 383 having pilot chambers 381, 382 positioned opposite to each other.
  • the delivery pressure of a fixed displacement hydraulic pump 385 is introduced to the pilot chamber 381 through a pilot line 384 and the maximum load pressure is introduced to the pilot chamber 382 through a pilot line 386, with a spring 387 disposed on the same side as the pilot chamber 382.
  • This arrangement enables to hold the delivery pressure of the hydraulic pump 385 higher than the maximum load pressure among a plurality of hydraulic actuators by a pressure value corresponding to the resilient strength of the spring 387.
  • Fig. 26 shows such a modification. More specifically, in Fig. 26, a hydraulic pump 390 is connected to a flow control valve 391 consisted of a main valve, a pilot valve and a pressure compensating valve which are combined as mentioned above, and produces a delivery flow rate adjusted by a pump flow control device 392. An unloading valve 393 is connected between the hydraulic pump 390 and the flow control valve 391, and the flow control valve 391 is associated with an operation device 394.
  • An operated signal from the operation device 394 is sent to a control device 395 which applies a control signal to a pilot valve driver part 396 of the flow control valve 391 for controlling the opening degree of the pilot valve.
  • the operated signal sent to the control device 395 is also applied to a processing device 397 which calculates a required flow rate of the flow control valve 391 from the map previously stored in a storage device 398, and then sends a calculated signal to the pump flow control device 392.
  • the processing device 397 calculates a setting pressure of the unloading valve 393 from another map previously stored in the storage device 398, and then sends a calculated signal to the unloading valve 393. This allows the delivery pressure of the hydraulic pump 390 to be controlled equal to a pressure obtained from the map previously stored in the storage device 398 as a function of the operated signal.
  • the differential pressure Ps - Pl max represented by the first term in the right side of the foregoing equation (1) cannot be controlled to be constant. Therefore, the pressure compensating function obtainable with the first term in the right side cannot be achieved. In the combined operation, however, that differential pressure remains common to all of the flow control valves associated with the respective hydraulic actuators, so the flow distributing function can still be achieved. Further, since the second and third terms in the right side of the equation (1) are not related to the pump delivery pressure Ps, the harmonizing function and/or the self-pressure compensating function based on the flow distributing function can be achieved in case of setting ⁇ , ⁇ to any values other than zero.
  • the pump control means may be associated with a simple relief valve for holding the delivery pressure of the hydraulic pump at constant.

Claims (12)

  1. Ein hydraulisches Antriebssystem, das umfaßt: wenigstens eine Hydraulikpumpe (1), wenigstens ein erstes und ein zweites Hydraulik-Betätigungselement (6, 7), die über entsprechende Hauptleitungen (2 - 3) mit der Hydraulikpumpe (1) verbunden sind und mittels von der Hydraulikpumpe (1) abgegebenen Hydraulikfluids angetrieben werden; eine erste und eine zweite Strömungssteuerung-Ventileinrichtung (8, 9), die zwischen der Hydraulikpumpe(1) und dem ersten bzw. dem zweiten Hydraulik-Betätigungselement (6, 7) mit den jeweiligen Hauptleitungen (2, 3) verbunden sind, eine Pumpsteuereinrichtung (10) zum Steuern eines Förderdrucks (Ps) der Hydraulikpumpe (1); wobei sowohl die erste als auch die zweite Strömungssteuerung-Ventileinrichtung (8, 9) versehen ist mit einer ersten Ventileinrichtung (15, 74), deren Öffnungsgrad in Abhängigkeit vom eingestellten Wert einer Bedienungseinrichtung (30) veränderlich ist, und einer zweiten Ventileinrichtung (16, 75), die mit der ersten Ventileinrichtung (15, 74) in Reihe geschaltet ist, um einen Differenzdruck (ΔPz) zwischen dem Einlaßdruck (Pz) und dem Auslaßdruck (Pl) der ersten Ventileinrichtung (15, 74) zu steuern; eine Steuereinrichtung (43-49, 51), die sowohl der ersten als auch der zweiten Strömungssteuerung-Ventileinrichtung (8, 9) zugeordnet ist, um die zweite Ventileinrichtung (16, 75) dazu zu veranlassen, den Differenzdruck (ΔPz) zwischen dem Einlaßdruck (Pz) und dem Auslaßdruck (Pl) der ersten Ventileinrichtung (15, 74) auf der Grundlage des Einlaßdrucks (Pz) und des Auslaßdrucks (Pl) der ersten Ventileinrichtung (15, 74), des Förderdrucks (Ps) der Hydraulikpumpe (1) und des maximalen Lastdrucks (Pl max) zwischen dem ersten und dem zweiten Hydraulik-Betätigungselement (6, 7) zu steuern,
    dadurch gekennzeichnet, daß
       sowohl die erste als auch die zweite Strömungssteuerung-Ventileinrichtung (8, 9) umfaßt: ein Hauptventil (11, 70) vom Typ eines Ventils mit Ventilsitz, mit einem Ventilkörper (21) zum steuern der Verbindung zwischen einem Einlaßkanal (17) und einem Auslaßkanal (18), die beide mit dem Hauptkreis (2-5) verbunden sind, einem variablen Begrenzer (23), der den Öffnungsgrad dieser Verbindung in Abhängigkeit von den Bewegungen des Ventilkörpers (21) ändern kann, und einer Gegendruckkammer (24), die über den variablen Begrenzer (23) mit dem Einlaßkanal (17) in Verbindung steht und einen Steuerdruck (Pc) erzeugt, um den Ventilkörper (21) in die Ventilschließrichtung zu zwingen; und einen Vorsteuerkreis (12-14, 71-73), der zwischen der Gegendruckkammer (24) und dem Auslaßkanal (18) des Hauptventils (11, 70) angeschlossen ist;
       die erste Ventileinrichtung (15, 74) in den Vorsteuerkreis (12-14, 71-73) als Vorsteuerventil (15, 74) geschaltet ist, um eine durch den Vorsteuerkreis sich bewegende Vorsteuerströmung zu steuern, und die zweite Ventileinrichtung (16, 75) in den Vorsteuerkreis als Hilfsventileinrichtung (16, 75) geschaltet ist, um einen Differenzdruck (ΔPz) zwischen dem Einlaßdruck (Pz) und dem Auslaßdruck (Pl) des Vorsteuerventils (15, 74) zu steuern; und
       die Steuereinrichtung (43-49, 51) die Hilfsventileinrichtung (16, 75) sowohl für die erste als auch für die zweite Strömungssteuerung-Ventileinrichtung (8, 9) steuert, derart, daß der Differenzdruck (ΔPz) zwischen dem Einlaßdruck (Pz) und dem Auslaßdruck (Pl) des Vorsteuerventils (15, 74) eine durch die folgende Gleichung ausgedrückte Beziehung zu einem Differenzdruck zwischen dem Förderdruck (Ps) der Hydraulikpumpe (1) und dem maximalen Lastdruck (Pl max) des ersten und des zweiten Hydraulik-Betätigungselementes (6, 7), zu einem Differenzdruck zwischen dem maximalen Lastdruck (Pl max) und dem Eigenlastdruck (Pl) eines jeden Hydraulik-Betätigungselementes (6, 7), und zu dem Eigenlastdruck (Pl) besitzt,

    ΔPz = α (Ps - Pl max) + β (Pl max - Pl) + γ Pl,
    Figure imgb0075

    wobei
    ΔPz:   Differenzdruck zwischen dem Einlaßdruck (Pz) und dem Auslaßdruck (Pl) des Vorsteuerventils (15, 74),
    Ps:   Förderdruck der Hydraulikpumpe (1),
    Pl max:   maximaler Lastdruck zwischen dem ersten und dem zweiten Hydraulik-Betätigungselement (6, 7),
    Pl:   Eigenlastdruck des ersten bzw. des zweiten Hydraulik-Betätigungselementes (6, 7),
    α,β,γ:   erste, zweite und dritte Konstanten,
    wobei die ersten, zweiten und dritten Konstanten α, β, γ auf entsprechende vorgegebene Werte gesetzt sind.
  2. Ein hydraulisches Antriebssystem gemäß Anspruch 1, dadurch gekennzeichnet, daß für die erste Konstante α die Beziehung α ≦ K gilt, wobei angenommen ist, daß K das Verhältnis (As/Ac) ist zwischen der Druckaufnahmefläche (As) des Ventilkörpers (21) des Hauptventils (11, 70), die über den Einlaßkanal (17) dem Förderdruck (Ps) der Hydraulikpumpe (1) unterworfen ist, und der Druckaufnahmefläche (Ac) des Ventilkörpers (21) des Hauptventils (11, 70), die dem Steuerdruck (Pc) der Gegendruckkammer (24) unterworfen ist.
  3. Ein hydraulisches Antriebssystem gemäß Anspruch 2, dadurch gekennzeichnet, daß die zweite und die dritte Konstante (β, γ) auf Null gesetzt sind.
  4. Ein hydraulisches Antriebssystem gemäß Anspruch 1, dadurch gekennzeichnet, daß die erste Konstante (α) auf einen beliebigen gewünschten, positiven Wert gesetzt ist, der dem proportionalen Verstärkungsfaktor der Hauptströmungsrate des Hauptventils (11, 70) in bezug auf den eingestellten Wert der Bedienungseinrichtung (30) entspricht.
  5. Ein hydraulisches Antriebssystem gemäß Anspruch 1, dadurch gekennzeichnet, daß die zweite Konstante (ß) auf einen beliebigen gewünschten Wert gesetzt ist, der auf der Harmonisierung des kombinierten Betriebs des zugehörigen Hydraulik-Betätigungselementes (6, 7) mit einem oder mehreren anderen Hydraulik-Betätigungselementen (7, 6) basiert.
  6. Ein hydraulisches Antriebssystem gemäß Anspruch 1, dadurch gekennzeichnet, daß die dritte Konstante (γ) auf einen beliebigen gewünschten Wert gesetzt ist, der auf den Betriebseigenschaften des zugehörigen Hydraulik-Betätigungselementes (6, 7) basiert.
  7. Ein hydraulisches Antriebssystem gemäß Anspruch 1, dadurch gekennzeichnet, daß die Steuereinrichtung mehrere in jedem der Hilfsventile (16, 75) vorgesehene Hydraulik-Steuerkammern (43-46) für die erste und die zweite Strömungssteuerung-Ventileinrichtung (8, 9) und ein Leitungsmittel (47-49, 51) umfaßt, um den Förderdruck (Ps) der Hydraulikpumpe (1), den maximalen Lastdruck (Pl max) und den Einlaßdruck (Pz) und den Auslaßdruck (Pl) des Vorsteuerventils (15, 74) direkt oder indirekt in die mehreren Hydraulik-Steuerkammern (43-46) einzuleiten, wobei die jeweiligen Druckaufnahmeflächen (as, al, az, am) der mehreren Hydraulik-Steuerkammern (43, 46) so bemessen sind, daß die erste, die zweite und die dritte Konstante (α, β, γ) gleich den jeweiligen vorgegebenen Werten sind.
  8. Ein hydraulisches Antriebssystem gemäß Anspruch 7, dadurch gekennzeichnet, daß das Hilfsventil (124, 125) zwischen der Gegendruckkammer (24) des Hauptventils (102, 103) und dem Vorsteuerventil (120, 121) angeordnet ist, die mehreren Hydraulik-Steuerkammern eine erste Hydraulik-Steuerkammer (131), die das Hilfsventil (124) in die Ventilöffnungsrichtung zwingt, und eine zweite, eine dritte und eine vierte Hydraulik-Steuerkammer (132-134), die das Hilfsventil (124) in die Ventilschließrichtung zwingen, umfassen und das Leitungsmittel eine erste Leitung (12, 135) zum Einleiten des in der Gegendruckkammer (24) des Hauptventils (102) herrschenden Steuerdrucks (Pc) in die erste Hydraulik-Steuerkammer (131), eine zweite Leitung (13) zum Einleiten des Einlaßdrucks (Pz) des Vorsteuerventils (120) in die zweite Hydraulik-Steuerkammer (132), eine dritte Leitung (136) zum Einleiten des maximalen Lastdrucks (Pl max) in die dritte Hydraulik-Steuerkammer (133) und eine vierte Leitung (137) zum Einleiten des Förderdrucks (Ps) der Hydraulikpumpe (1) in die vierte Hydraulik-Steuerkammer (134) umfaßt.
  9. Ein hydraulisches Antriebssystem gemäß Anspruch 8, dadurch gekennzeichnet, daß sowohl die erste als auch die zweite Strömungssteuerung-Ventileinrichtung (270) dadurch aufgebaut wird, daß das Hauptventil (271) und das Hilfsventil (272) in eine einteilige Struktur eingebaut sind.
  10. Ein hydraulisches Antriebssystem gemäß Anspruch 1, dadurch gekennzeichnet, daß die Steuereinrichtung versehen ist mit elektromagnetischen Funktionselementen (202A, 202B), die in jeder der Hilfsventileinrichtungen (202, 203) für das erste und für das zweite Strömungssteuerungsventil (200, 201) vorgesehen sind, Druckerfassungseinrichtungen (204-209) zur direkten oder indirekten Erfassung des Förderdrucks der Hydraulikpumpe (1), des maximalen Lastdrucks und des Einlaßdrucks und des Auslaßdrucks des Vorsteuerventils (15, 74) und einer Verarbeitungseinrichtung (213) zum Berechnen eines Differenzdrucks zwischen dem Einlaßdruck und dem Auslaßdruck des Vorsteuerventils auf der Grundlage der von den Druckerfassungseinrichtungen erfaßten Signale und zum anschließenden Ausgeben eines berechneten Differenzdrucksignals an die elektromagnetischen Funktionselemente der Hilfsventileinrichtungen, wobei die erste, die zweite und die dritte Konstante α, β, γ in der Verarbeitungseinrichtung als die entsprechenden vorgegebenen Werte im voraus eingestellt werden.
  11. Ein hydraulisches Antriebssystem gemäß Anspruch 1, dadurch gekennzeichnet, daß die Pumpsteuereinrichtung ein Pumpenregulator (10) vom Lasterfassungstyp ist, um den Förderdruck der Hydraulikpumpe (1) um einen vorgegebenen Wert über dem maximalen Lastdruck zwischen dem ersten und dem zweiten Hydraulik-Betätigungselement (6, 7) zu halten.
  12. Ein Bagger, der das hydraulische Antriebssystem gemäß einem der Ansprüche 1 bis 11 enthält, wobei die Betätigungselemente dazu vorgesehen sind, die Mehrzahl von Funktionsorganen, die einen Schwenkkörper, einen Auslegerarm und einen Becher enthalten, anzutreiben.
EP19880201351 1987-06-30 1988-06-29 Hydraulisches Antriebssystem Expired - Lifetime EP0297682B1 (de)

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JP16270387 1987-06-30
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US4945723A (en) 1990-08-07
CN1011526B (zh) 1991-02-06
EP0297682A2 (de) 1989-01-04
AU603907B2 (en) 1990-11-29
CN1031270A (zh) 1989-02-22
DE3876518T2 (de) 1993-05-06
EP0297682A3 (en) 1989-04-12
KR920007653B1 (en) 1992-09-14
IN171522B (de) 1992-11-07
AU1842688A (en) 1989-01-05
KR890000799A (ko) 1989-03-16

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